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University of Wollongong University of Wollongong
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University of Wollongong Thesis Collection 1954-2016 University of Wollongong Thesis Collections
1974
An investigation into the burning rate of liquid petroleum gas in a multi-An investigation into the burning rate of liquid petroleum gas in a multi-
cylinder engine cylinder engine
M. G. Byrne Wollongong University College
Follow this and additional works at: https://ro.uow.edu.au/theses
University of Wollongong University of Wollongong
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Recommended Citation Recommended Citation Byrne, M. G., An investigation into the burning rate of liquid petroleum gas in a multi-cylinder engine, thesis, School of Civil, Mechanical and Mining Engineering, University of Wollongong, 1974. https://ro.uow.edu.au/theses/2925
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AN INVESTIGATION INTO THE BURNING RATE OF LIQUID
PETROLEUM GAS IN A MULTI-CYLINDER ENGINE.
M.G.BYRNE
B.E. N.S.W., Stud.I.E.Ausi.
Thesis submitted as part requirement for the
Degree of Master of Engineering Science.
School of Civil, Mechanical and Mining Engineering.
Wollongong University College.
University of New South Wales.
December 1974
IUNiVERSiTY OF WOLLONGONG
L.;-JR ARY
S’OSO
SUMMARY
The burning rates of liquid petroleum gas and petrol fuels in a
multicylinder spark ignition engine are examined and compared.
Findings are based on the pressure-time traces of cylinder
pres sure development during combustion,
Pressure tappings were fitted to three of the combustion
chambers on the test engine- a Holden six cylinder engine -
and a capacitance type pressure transducer together with
specialised electronic equipment were used for recording pressure
development.This made it possible to depict the pressure-time
traces for a particular cylinder on an oscilloscope screen for
observation and photographic recording.
The test engine, directly coupled to a dynamometer, was run on
both fuels under similar conditions and a comparison of the
pressure-time traces is made to establish differences in burning
rates. Operating conditions that were varied for the tests included
fuel type, ignition timing, air-fuel ratio, and engine speed.
Spark plug heat range and gap were not altered.
Results of earlier work conducted to compare the performance of
the engine for the two fuels, and which led to the present work,
are discussed briefly.
It was found from the investigation that liquid petroleum gas
exhibits a higher rate of pressure rise during combustion and
hence has a shorter combined ignition-burning period than that
of petrol. To realise full power from spark ignition engines
converted to liquid petroleum gas operation, the ignition timing
is normally revised to compensate for the differences in the
combined ignition-burning periods.
Variation of such parameters as ignition timing, air-fuel ratio
and engine speed resulted in similar characteristics being noted
for both fuels.
ACKNOWLEDGEMENT
The writer wishes to express his sincere thanks to Professor
S.E.Bonamy, Chairman, Department of Mechanical Engineering for
the opportunity of working in the Department and for his
assistance and guidance in the preparation of this report.
CONTENTS
SUMMARY
ACKNOWLEDGMENTSPage
SECTION I INTRODUCTION 1
SECTION II TEST EQUIPMENT AND EXPERIMENTAL PROCEDURE
(a) Engine and Dynamometer. 3
(b) Operating Equipment for Petrol. 4
(c) Operating Equipment for Propane. 5
(d) Air Flow Measuring Equipment. 6
(e) Equipment for Cylinder Pressure Measurement
and Recording. 6
SECTION III EXPERIMENTAL RESULTS AND DISCUSSION
General. 10
(a) Ignition-Burning Periods and Pressure
Development. 13
(b) Effect of Engine Speed. 15
(c) Effect of Ignition Timing. l6
(d) Effect of Air Fuel Ratio. 19
SECTION IV CONCLUSIONS AND BIBLIOGRAPHY
Conclusions. 21
Bibliography. 24
SECTION V APPENDICES
A. Equations for Engine Performance Factors. 27
SECTION V (cont*d)
B. Orifice Plate Calculations Using
B.S. 1042 Part I. 29
C. Tabulation of Test Data 34
SECTION I
INTRODUCTION
INTRODUCTION
Although liquified petroleum gas (L.P.G.) offers some advantages
over petrol in pollution control and considerable reserves of
L.P.G. are available in Australia, the conversion of motor vehicles
to L.P.G. operation has been limited by insufficient economic1incentives. A recent report issued by the Bureau of Transport
Economics estimates that known Australian reserves of L.P.G.
would be sufficient for 80 years supply at the present extraction
rate and that 14$ of Australian motor vehicles could be powered
by the supply. This represents a valuable energy source that is
currently being exported to other countries.
It is generally agreed upon by those involved with the conversion
of spark ignition engines from petrol to L.P.G. operation, that
it is necessary to revise the ignition timing to realise full
power from the engine. This revised ignition timing required by
the L.P.G. indicates that there are differences in the burning
rates of the two fuels.
Earlier tests prompted by the increasing awareness of air pollution
caused by the automobile engine and the introduction of laws2,3to control exhaust emissions, and carried out by the author ,
also indicated these differences. The tests, the results of
which will be discussed in Section III, were carried out to
compare the performance of a multicylinder engine using petrol
and L.P.G. as fuels. Results showed that, for the engine concerned,
L.P.G. required the ignition timing to he retarded by between
2 and 4 degrees from that required for petrol operation. Since
the peak cylinder pressures for the two fuels should occur at
approximately the same crank angle position at a given engine
speed for maximum torque, this difference in timing indicated
a faster combined ignition-burning period for the L.P.G.
The present work is a continuation of the earlier comparisons
and concerns an investigation into the differences in burning
rates between L.P.G. and petrol fuels in the same multicylinder
engine.
The rate of cylinder pressure rise has been shown to be strongly4 5 6 7dependent on the combustion rate ’ 9 9 and hence the findings
presented here are based on pressure-time traces recorded for the
two fuels at various operating conditions. Energy liberation
in the combustion process is dependent on the flame front area,
the concentration of the unburned mixture at the flame front, and
the rate of reaction relative to the unburned portion. Variations
in these parameters would also result in similar changes in the
rate of pressure rise.
For convenience, the liquid petroleum gas is referred to here as
Propane, the major constituent of the fuel.
SECTION II
TEST EQUIPMENT AND
EXPERIMENTAL PROCEDURE
FIGURE 1 s GENERAL VIEW OF TEST ENGINE AND EQUIPMENT
- 3 -
TEST EQUIPMENT AND
EXPERIMENTAL PROCEDURE
(a) ENGINE AND DYNAMOMETER
A general view of the test equipment is shown in Figure 1. Thet itest engine was a six cylinder, four stroke Holden engine of
3.062 inch bore and 3.125 inch stroke, having a capacity of
138 cubic inches.
Several modifications were made to the engine. The compression
ratio was increased from 6.5:1 to 8.0:1 to make it more comparable
with present standards and more suitable for both fuels. The
cylinder head face was machined and the volume of the chambers
measured by filling with light oil from a burette to ensure that
the required ratio was obtained. The vacuum advance mechanism on» »the Bosch distributor was disconnected and the distributor
clamp was modified to allow easy timing adjustment over a wide
range.
To measure ignition advance, a coupling on the engine-dynamometer
shaft was graduated in crank degrees for a range of 90 degrees
and a fixed pointer set at zero degrees to correspond with top
dead centre of number 1 cylinder. The top dead centre position
was obtained using a dial indicator. A Dawe stroboflood, triggered
by an induced current from the high tension lead of number 1
cylinder, illuminated the scale on each firing of the cylinder
enabling the ignition advance angle to be read.
- 4 -
The cooling system for the engine consisted of the normal radiator,. t ihelt driven water pump and fan with additional make-up water
being supplied through the radiator drain plug and withdrawn
through the radiator overflow tube. The temperature of the coolant
water was measured by a mercury thermometer fitted into the radiator
header tank at the point where the water returns from the engine.
The coolant temperature was maintained at 80°C during the tests.
The engine was coupled by a drive shaft to a Heenan and Froudet tDynamatic dynamometer Mark I G.V.A. having a capacity of 150 BHP
at 2400/6000 rpm. The water cooled dynamometer was equipped with
speed control which allowed the speed to be set and kept constant
for varying engine loadings. The calibration of the engine tachometer
on the dynamometer control desk was checked with a Smiths hand tachometer.
(b) OPERATING EQUIPMENT FOR PETROL.
A single barrel downdraught Stromberg BX0V.1 carburettor was
used for petrol operation. To allow variation of the air fuel
ratio, the main jet was drilled oversize and an adjustable
tapered needle was inserted to control the fuel flow rate.
Premium grade petrol was used and was supplied to the carburettor
by the normal AC mechanical fuel pump. Petrol consumption rates were
metered by the positive displacement method. A two way valve
was used to switch from the supply tank to a pipette with a
graduated volume of 100 cc*s.
FIGURE 2 i CONVERTOR AND CARBURETTOR FOR LPG OPERATION
- 5 -
(c) OPERATING EQUIPMENT FOR PROPANE
For propane an Inrpco model 100-8 carburettor vas used together
with an Inrpco model JB convertor (Figure 2) which converted the
propane from its high pressure liquid phase to a vapour at the
low pressure required at the carburettor. Heat for vapourising
the propane was supplied by hot water from the cooling system;t tthe inlet temperature being controlled to prevent freezing
of the unit.
The propane was supplied from 100 lb liquid withdrawal bottles
and from 20 lb bottles which were inverted to ensure liquid
withdrawal. The liquid petroleum gas, referred to as propane in
this report, was stated by the suppliers to have the following
approximate composition: 88$ propane, 2$ butane and 10$ propylene.
The mass flow rate to the engine was measured by switching from
the 100 lb bottle to a 20 lb bottle for a period of 30 or 60
seconds. The 20 lb bottle was weighed carefully on a set of
Wedderburn precision scales before and after the chosen time
interval.
FIGURE 3
X X
: PLAN VIEW OF COMBUSTION CHAMBER SHOWING LOCATION OF PRESSURE TAPPING POINTS.
- 6 -
(d) AIR FLOW MEASURING EQUIPMENT
To measure the air flow rate to the engine the intake was connected
by a short length of smooth bore, heavy duty rubber hose of 2^ inch
inside diameter to a large surge tank which damped out pulsations
prior to metering. Atmospheric air was drawn into the tank through
a short pipe containing a metering orifice plate designed to the
specifications of B.S. 1042. (Appendix 2) The differential
pressure across the orifice plate was measured with an Askania
micromanometer model 6-0042/72.
(e) EQUIPMENT FOR CYLINDER PRESSURE MEASUREMENT AND RECORDING.
To allow the cylinder pressure to be recorded, tappings were
made into three of the six combustion chambers (cylinders 2, 4
and 6), the direction of entry being from the top of the cylinder
head through the water jackets. The point of entry of the tappings
into the roof of the combustion chambers was opposite the spark
plug location as shown in Figure 3»
To minimise errors in pressure measurement, the length of the
passage between the combustion chamber and the pressure transducer
was kept to a minimum (approx. 4 inches for each tapping) and
copper washers were used to ensure that no additional clearance
volumes resulted when the transducer was fitted to the tapping
point. The passage hole diameter was 0.047 inches.
PressureTransducer
MR 313Switching Panel
MR 220F FM Gauge
Amplifier
MR 209D Lower /Driver Vertical^
Amplifier Beam \(
Tektronix Type 563
Dual BeamicilliscopiandCamera
Horizontal Beams Synchronised Trigger
Upper Vertical Beam
M 738 Photo-cell
andDegree Marker
Unit
coupled to engine driveshaft
MR 281Marker Amplifier
FIGURE 4 BLOCK DIAGRAM OF CYLINDER PRESSURE MEASURING AND RECORDING EQUIPMENT WITH TYPICALEXAMPLE OF A PHOTOGRAPHIC RECORD OF A SINGLE PRESSURE-TIME TRACE
- 7 -
A second tapping was fitted to cylinder 6 to allow possible
differences in pressure recordings from the two tappings to be
examined. The tapping was made horizontally from the back of
the head entering the side wall of the combustion chamber near the
exhaust valve which resulted in a shorter passage length of approx,
two inches. It was later found that the pressure-time traces
from this tapping were somewhat erratic, possibly due to the
entry position of the tapping in the combustion chamber.
The pressure sensing and recording equipment is represented diagramatically in Figure 4.
The pressure transducer used was a water-cooled Minirack type
G202 capacitance unit. The output of the pressure transducer was
fed via the coaxial cable through the Minirack MR220F FM gauge
amplifier and the MR313 switching panel to a MR209D driver
amplifier. The output from the driver amplifier was fed to the
lower beam of a Tektronix Type 565 dual beam oscilliscope fitted
with Type 3A75 amplifiers on both the lower and upper vertical
beam inputs.
The output from the photocell unit was fed via a MR292B photocell
pre-amplifier to the horizontal beam trigger circuit of the
oscilliscope. The two horizontal sweeps of the oscilliscope were
locked together to achieve the synchronisation between the
pressure-time trace and the degree marker points.
- 8 -
The output from the degree marker disc and magnetic pick-up
was fed via a MR281 marker amplifier to the upper vertical beam
of the dual beam oscilliscope.
The results obtained on the oscilliscope were recorded with the
special Polaroid camera attachment for the oscilliscope. Controls
on the oscilliscope allowed a single trace to be displayed on the
screen when required. The camera shutter was held open for a
selected period during which the single trace was manually triggered.
A typical example of a single pressure-time trace with degree
marker points is shown in Figure 4, The centre point of the group
of five degree marker points represents top dead centre. These
five points are spaced at intervals of 20 crank angle degrees,
while the outer points are spaced at 40 degree intervals. On
the vertical scale, the distance between grid lines represents
60 psi pressure.
The particular combination of the electrical equipment used did
not allow the static calibration of the pressure transducer and
equipment against a dead-weight tester to be made. The equipment
was dynamically calibrated against a bourden-tube type pressure
gauge fitted with a Schrader non-return valve. The gauge and
transducer were fitted to the number 6 cylinder and the engine
was operated at various conditions to give a range of peak
pressures for calibration. The pressure gauge was statically
checked against a dead-weight tester.
- 9 -
Although the accuracy of this method might not have matched that
of the direct static calibration of the transducer and equipment,
it was considered satisfactory as the test results were to he
analysed on a conqmrative basis. Calibration of the transducer
showed the response to be essentially linear and, for the equipment
settings used for the tests, gave a scale of 60 psi per major
vertical division on the oscilliscope screen.
SECTION III
EXPERIMENTAL RESULTS
AND DISCUSSION
Volumetrie E
fficiency
(70)
Brake
Horse
Power
60
50
40
30
20
10
o▲—PetrProp
olane f
s---k
;6
rJ».. .......i: k*. '< •« >
••
»»<
////
v ~k--- *i r*7--
% 1
1 %» < %\
[!\ .c
/ // /
/k * '• »% %
‘i>k _
____
oV\
• • è 9 •bhpToreque
105
95
85
75
1000 2000 3000 4000
Engine Speed (rpm)
FIGURE 5 : PERFORMANCE CURVES.
90
85
80
75
70
651000 2000 3000 4000
Engine Speed (rpm)
o▲—PetrProp
olane
o»K\ N
▲\
V. >\\\xi----ik--—<iT \ V N
'i
>\V ,\ i
Vi►L
FIGURE 6 : VOLUMETRIC EFFICIENCY
Torque (ft.lbs)
35COO)QJUbD<L>'TDO)oGCO>rOGO•H4->GW)
30
25
20
C)/ _///
/
o- \
\*c/l‘1 fl
/► ---i/
/
L A i
II¿11A
oA.
PetrolPropane
0 1000 2000 3000 4000Engine Speed (rpm)
FIGURE 7 : IGNITION ADVANCE FOR OPTIMUM OUTPUT AT FULL LOAD.
10 -
EXPERIMENTAL RESULTS
AND DISCUSSION
By way of introduction to the discussion of experimental results,
three figures showing results of earlier performance tests,
carried out on the same engine * , are included.
Figures 5 and 6 show the performance curves and corresponding
volumetric efficiencies for the two fuels over the speed range
800-4000 rpm. Conversion to propane resulted in an average decrease
in performance of slightly more than 5$ over the speed range. This
was shown to he attributable mainly to the drop in volumetric
efficiency (Fig.6) of the engine when converted to propane operation.
The decrease in the volumetric efficiency resulted basically
from the displacement of air entering the carburettor by the
propane vapour and from the design characteristics of the particular
carburettor used.
Figure 7 shows the ignition requirements of the fuels for full
load optimum performance* As mentioned earlier, the differences
in ignition timing requirements( approx. 2° retard up to 1500rpm
and 3° retard above 1500 rpm for propane ) indicate a faster
combined ignition-burning period for propane, since the peak
cylinder pressures for the two fuels should occur at approximately
the same crank angle for a given engine speed for maximum torque.
The present experimental results are used to examine the combined
ignition-burning periods of the two fuels and the effects of
variations in ignition timing, engine speed and air-fuel ratio
360
b0COCl
O)5 -iPCOCOa)cu
300
240
180
120
60
0160 120 80 40 TDC 40 80 120
Time, degrees of crankshaft rotation
FIGURE 8 : TYPICAL PRESSURE-TIME TRACES SHOWING CYLINDER-TO-CYLINDER VARIATION. CYLINDERS 2, 4, 6
FIGURE 9 : PRESSURE-TIME TRACES SHOWING TYPICAL CYCLE-BY-CYCLE VARIATION. CYLINDER 6.
- l i
on the pressure-time curves.
To conduct such a comparison using the pressure-time curves of
the two fuels, some point on the curve signifying the completion of combustion must he chosen.
7 8 QReports by Rassweiler et al and Marvin et al on the
examination of flame movement and pressure development using
photographic techniques to record flame front movement indicate
that, for higher compression engines, the completion of
combustion occurs very close to the point of maximum pressure.
Variations of the air-fuel ratios and ignition timing away from
the optimums reduced the amount of combustion completed at the peak pressure.
Cylinder pressure variation is a fundamental problem in spark
ignition engines. Patterson^ and many others have shown that no
two cylinders of a multicylinder engine produce identical average
pressure records. For this reason pressure-time traces were
recorded for each of the three cylinder tappings. Typical
cylinder-to-cylinder variations are illustrated in Figure 8.
Because of the variation between cylinders, it was decided to
use one particular cylinder (cylinder 6) for the test program.
Cycle-by-cycle variability for a given cylinder can also present
problems, particularly where the pressure-time curves are recorded
- 12 -
by photographing single traces on the oscilloscope screen. An
advantage of having the Polaroid camera was that each photograph
of a pressure-time trace could be examined during the tests and
judged as to its suitability as an *averag&* trace. Figure 9
illustrates the cycle-by-cycle variability recorded over several
consecutive cycles. Although it was observed that, for a particular
set of operating conditions, cycle-by-cycle variability gave a
range of peak pressures, the crank angle position at which the
peak pressure occurred remained essentially constant. Thus the
point at which peak pressure occurred was used to signify the
completion of combustion for the purpose of comparison of ignition
-burning periods for optimum pressure development conditions.
Discussion of the experimental work and results is divided into
four sections to enable comparisons to be made between the two
fuels in relation to the following-
(a) Ignition-burning periods and pressure development.
(b) Effects of engine speed.
(c) Effects of ignition timing.
(d) Effects of air-fuel ratio.
Pressure (
psig)
Pressure (
psig)
420'
360300
240
180
120
60
160 120 80 40 TDC 40 80 120Time, degrees of crankshaft rotation.
FIGURE 10 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AND PROPANE AT 2000 RPM, OPTIMUM OUTPUT. IGNITION ADVANCE-PETROL 32°BTDC, PROPANE 28°BTDC
160 120 80 40 TDC 40 80 120 160Time, degrees of crankshaft rotation.
FIGURE 11 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AND PROPANE AT 2500 RPM, OPTIMUM OUTPUT. IGNITION ADVANCE-PETROL 32°BTDC, PROPANE 29°BTDC
360
300
160 80 TDC 80 160Time, degrees of crankshaft rotation.
FIGURE 12 : TYPICAL PRESSURE-TIME TRACES FOR PETROLAND PROPANE AT 3000 RPM, OPTIMUM OUTPUT. IGNITION
ADVANCE - PETROL 32°BTDC, PROPANE 29°BTDC.
- 13 -
(a) IGNITION-BURNING PEROIDS AND PRESSURE DEVELOPMENT.
To compare the combined ignition-burning peroids of the two
fuels, full throttle tests were conducted at a number of engine
speeds and at optimum ignition and mixture settings.(ie, leanest
mixture for best torque and minimum advance for best torque.)
For reasons mentioned earlier, the point at which peak pressure
occurred is chosen to signify the completion of combustion for
the optimum conditions.
Figures 10, 11 and 12 illustrate typical pressure-time curves
for speeds of 2000, 2500 and 3000 rpm respectively. The ignition
timings used are noted on each trace and combined ignition
burning periods and maximum rates of pressure rise for the figures are given in Table 1.
From the pressure-time curves it is observed that propane has
a shorter combined ignition-burning period. For example, at
2500 rpm (Fig.ll) the period for propane was scaled at 64 crank
angle degrees (4.26 msec) compared with 72 crank angle degrees
(4.80 msec) for petrol; a difference of 8 degrees (0.54 msec).
Peak pressure is seen to occur earlier for the propane fuel for
optimum output conditions. For example, seperation of the peaks
at 2500 rpm (Fig.11) is approximately 5 degrees (0.33 msec).
The rate of pressure increase on the pressure-time curves is
- 14 -
strongly dependent on the combustion rate. From the maximum
gradients measured from the pressure-time curves it is found
that the propane fuel generally acheived a higher rate of
pressure rise, ie, a faster combustion rate. For example, from
Figures 10, 11 and 12 the maximum rates of pressure rise are
72 psi/msec, 90 psi/msec and 144 psi/msec respectively for petrol
as compared to 108 psi/msec, 135 psi/msec and 144 psi/msec
respectively for propane.
When comparing the pressure-time traces for the two fuels, the
differences in pressure development are usually evident before
top dead centre is reached. This would indicate a slight increase
in the negative work (compression) occurring with the propane.
Throughout the tests peak pressures for propane were found to be
greater than those for petrol at corresponding operating conditions
even though the engine output was less with the propane. The
greater rate of pressure rise appears to result in the higher
peak pressures. However, because of this steeper rate of pressure
increase, the area under the propane pressure-time curve after
top dead centre tends to be slightly less than that for petrol.
As the difference between this positive area and the negative
(compression) work area (found to be slightly greater for propane)
is proportional to engine output, the smaller resultant area
would agree with the lower engine outputs obtained with propane.
Table 1 gives torque values corresponding to the pressure-time
curves of Figures 10, 11 and 12.
Pressure (
psig)
Pressure (
psig)
420
360 -------
300 -------
240 -------
180-------
1 2 0 --------------
6 0 -------
160 120 80 40 TDC 40 80 120Time, degrees of crankshaft rotation.
FIGURE 13 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT SPEEDS OF 2000, 2500, AND 3000 RPM. IGNITION ADVANCE 32°BTDC.
Time, degrees of crankshaft rotation.
FIGURE 14 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT SPEEDS OF 2000, 2500 AND 3000 RPM. IGNITION ADVANCE - 2000 RPM 28°BTDC, 2500 AND 3000 RPM 29°BTDC.
FIGURE 15 : TYPICAL PRESSURE-TIME TRACE FOR PETROL AT SPEEDS OF 2000, 2500 AND 3000 RPM . IGNITION ADVANCE 40°BTDC.
FIGURE 16 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT SPEEDS OF 2000, 2500 AND 3000 RPM. IGNITION ADVANCE 40°BTDC.
- 15 -
(b) EFFECT OF ENGINE SPEED.
5 8As noted in publications by Litchy , Marvin and Best and
Rassweiler et al^ it bas generally been established that flame
speed increases with engine speed. The increase of turbulence of
the fuel mixture with engine speed and its effects on flame
growth and flame front velocity are considered the main reasons
for the increase. Ignition timing is normally advanced as engine
speed increases to maintain the desired pressure development
and engine output.
Pressure-time traces were recorded at selected engine speeds and
ignition timings with full throttle; the air-fuel ratio being
adjusted to the leanest mixture for best torque.
The effects of engine speed on pressure development for the two
fuels are illustrated in Figures 13 > 14, 15 and 16. Figures 13
and 14 are typical pressure-time traces for optimum ignition
timing at speeds of 2000, 2500 and 3000 rpm. Figures 15 and 16
are typical pressure-time traces for the same speeds but at a
constant ignition timing of 40° BTDC for both fuels.
It can be seen that both fuels exhibit similar characteristics
for the engine speed variation. As the engine speed increases
the peak pressures are reduced in magnitude and the peak occurs. . . 9at a later crank angle position. Marvin, Wharton and Roeder
found that the flame velocity and the time rate of pressure
development increase only a little more slowly than engine speed.
This slight difference would explain the above characteristics.
420
360
160 120 80 40 TDC 40 80 120 160
Time, degrees of crankshaft rotation
FIGURE 17 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT 2000 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
420
360^ 300bO£ 240
u 180Dco
3 120ueu60
160 120 80 40 TDC 40 80 120
Time, degrees of crankshaft rotation.
FIGURE 18 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT 2000 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
Pressure (
psig)
Pressure (
psig)
Time, degrees of crankshaft rotation.
FIGURE 19 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT 2500 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
420
360
300
240
180
120
60
160 120 80 40 TDC 40 80 120 160Time, degrees of crankshaft rotation.
FIGURE 20 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT 2500 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
420
360
300
a 2400)£ 180cocov 120uPM
60
160 120 80 40 TDC 40 80 120 160
Time, degrees of crankshaft rotation.
FIGURE 21 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT 3000 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
Time, degrees of crankshaft rotation.
FIGURE 22 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT 3000 RPM AND VARIOUS IGNITION ADVANCE SETTINGS.
100
90CO
rOr—'-U M—I80CDPcr u o E-170
60
/ / A
/ /
' l l - * -
— O — ._ > ̂
/)//
//
/ /
Y
---*----
///
/
o Petrol A Propane
0 10 20 30 40 50
Ignition advance (degrees)
FIGURE 23 : TORQUE VERSUS IGNITION ADVANCE FORPETROL AND PROPANE AT 2000 RPM.
100
CO
4->CM0)pcroH
90
80
70
60
//
/ s
s ' '
~~ — — *
O ™ IX«
^ XX
X .\\N
___________________
'N \ XX
7t//
4'&
o Petrol A Propane
0 10 20 30 40 50Ignition advance (degrees)
FIGURE 24 : TORQUE VERSUS IGNITION ADVANCE FORPETROL AND PROPANE AT 2500 RPM
100
CO,a-p
a>pcrPoE-<
90
80
70
60
<
- o — ̂»
/
/ s
/ // /
— h r
\<
3
////
» /(
V k
"77“< / /i
° Petrol * Propane
0 10 20 30 40 50
Ignition advance (degrees)
FIGURE 25 : TORQUE VERSUS IGNITION ADVANCE FORPETROL AND PROPANE AT 3000 RPM.
16 -
(c) EFFECT OF IGNITION TIMING.
The timing of ignition is an important factor in the pressure
development and power output of the spark ignition engine and
the requirements vary with fuel type and engine characteristics.
The optimum performance timing requirements for the test engine,
as found in earlier tests ’ , have already been shown in Figure 7*
During these earlier tests it was noted that the propane was
more sensitive to changes in ignition timing.
To examine the effects of the variation of ignition timing on
pressure development and engine performance at full throttle,
tests were conducted at selected engine speeds and with the air-
fuel ratio adjusted to leanest mixture for best torque. For each
engine speed the ignition timing was varied in steps from 2° BTDC
to 50° BTDC. Pressure-time traces and engine output were recorded
for each ignition timing, the test results being Illustrated in
Figures 17 through 25 and tabulated in Tables 2, 3 and 4.
The pressure-time curves obtained indicate that the fuels exhibit
similar characteristics with the variation of ignition timing.
As the point of ignition is advanced from 2° BTDC to 50° BTDC the
peak pressures increase for both fuels; the propane developing
- 17 -
slightly higher peak pressures because of the shorter combined
ignition-combustion period for the fuel.
The negative (compression) work area under the pressure-time
curves increases with the advance of ignition. As seen from the
typical pressure-time curves illustrated in Figures 17 through
22 the rate of increase of this negative work area with ignition
advance is greater with propane than with petrol, again as a
result of the shorter combined ignition-combustion period of propane.
Considering the differences between the combined ignition-
combustion periods of the two fuels, it can be reasoned that
for ignition timings further advanced from the optimum, the
engine output would be more sensitive with propane; whereas for
ignition timings retardedfrom the optimum, the engine output
would be more sensitive with petrol.
Actual results obtained are in agreement with the above reasoning.
Figures 23, 24 and 25 illustrate the engine torque versus ignition
timing for speeds of 2000, 2500 and 300Q rpm and it is seen from
these figures that engine output is more sensitive with the
propane for timing in advance of the optimum and more sensitive
with petrol for timing retarded from the optimum.
18 -
It was also found during the tests that, for the more retarded
ignition timings (10 and 2 BTDC), * late* pressure peaks occurred.
Typical examples are included in Figures 18, 21 and 22. These
*late* pressure peaks were more prevalent with propane at the
lower speeds hut occurred with similar frequency for both fuels
at the higher speeds, eg. 3000 rpm. The late ignition and
* delayed * combustion would be mainly responsible for these late peaks.
Pressure (
psig)
Pressure (
psig)
Time, degrees of crankshaft rotation,
FIGURE 26 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT 2000 RPM AND VARIOUS AIR FUEL RATIOS.
Time, degrees of crankshaft rotation.
FIGURE 27 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT 2000 RPM AND VARIOUS AIR FUEL RATIOS.
Pressu
re (
psig)
Pressure (
psig)
Time, degrees of crankshaft rotation,
FIGURE 28 : TYPICAL PRESSURE-TIME TRACES FOR PETROL AT 2500 RPM AND VARIOUS AIR FUEL RATIOS.
Time, degrees of crankshaft rotation,
FIGURE 29 : TYPICAL PRESSURE-TIME TRACES FOR PROPANE AT 2500 RPM AND VARIOUS AIR FUEL RATIOS.
100
90
_ 80 CO
rO r—I
S 7 0(U
ST 60OH50
40
0.7 0, 8 0.9 1.0 1.1 1.2 1.3 1.4 1.5Equivalence Ratio
---- 1----------- 1---------- 1-----------1-----------1----------- 1
11 13 15 17 19 21Air Fuel Ratio Petrol
I-----------1---------- 1---------- 1----------- 1----------- 1----------- 111 13 15 17 19 21 23Air Fuel Ratio Propane
FIGURE 30 : AIR FUEL RATIO VERSUS TORQUE FOR PETROL AND
0 . 6I—9
=0=^=0' x N
 â s xN<\ X \ \
AN\
\ * \\ 1 \ \o\\
\\\
\O \\
O▲
PetroPropa
Lle
\ '\\\
PROPANE AT 2000 RPM.
100
COrû4-1
M -l
<DacrnoE-*
90
80
70
60
50
40
0.6 0,7 0,8 0,9 1.0 1.1 1.2 1.3 1.4 1.5Equivalence Ratio
1-------9
----------------- »--------11
-------------- r -------13
--------------- 1------------------15
------- 1-------------------------1— ----------------------»17 19 21
Air Fuel Ratio Petroli ï
113 15 17 19 21 23
Air Fuel Ratio Propane.
8° — o— o*
4-4— 4 \
\ N\o\\», \
VA\\
VTo\
o▲
Petro!Propaiîe
\\
FIGURE 31 : AIR FUEL RATIO VERSUS TORQUE FOR PETROL AND PROPANE AT 2500 RPM.
- 19 -
(d) EFFECT OF AIR-FUEL RATIO.
The effects of air-fuel ratio on pressure development were
examined at various engine speeds. The tests were conducted at
full throttle with ignition timimg set for maximum engine output
at the optimum air-fuel ratio.(ie, leanest mixture for best torque)
Both fuels were found to exhibit similar characteristics with
variation of the air fuel ratio. Figures 26 and 27 (2000 rpm)
and 28 and 29 (2500 rpm) illustrate pressure-time curves typical
of those obtained during the tests. The corresponding torque
versus air fuel ratio relationships are shown in Figures 30 and
31 and tabulated in Tables 5 and 6. The equivalence ratio (based
°n 14.9 stoichiometric for petrol and 15.7 stoichiometric for
propane) is included to allow a more meaningful comparison of
performance versus * leanness * between the two fuels.
Conqmrison of the pressure-time curves with the corresponding
torque curves shows that for both fuels the maximum rates of
pressure rise and peak pressures occur for the air fuel ratios in
the region of the leanest mixture for best torque. As the air
fuel mixture is weakened from this setting, the rate of pressure
rise decreases and the peak pressure decreases and occurs at a later angle.
- 20 -
t t
Mixtures richer than those of the optimum region were obtained
with the petrol carburettor. However, as mentioned earlier, the
design of the propane carburettor prevented the richer mixtures from being achieved.
The pressure-time curves obtained for petrol indicate that rich
mixtures result in characteristics similar to those of weak
mixtures, ie, the rate of pressure rise decreases, and the peak
pressure decreases and occurs at a later angle. Marvin andg
Best studied flame speed by photographic techniques and found
that a decrease in flame speed was caused by using a mixture
richer or leaner than that giving maximum power. Similarly
Eassweiler, Witwrow and Cornelius studied combustion rates by
photographic techniques and found that a weakening of the
mixture resulted in an increased combustion time acconqmnied by
the decrease in pressure rise rates and the lower peak pressures occurring at later angles.
SECTION IV
CONCLUSIONS AND BIBLIOGRAPHY
21
CONCLUSIONS
The test program reported above leads to the following conclusions;
The LPG fuel has a shorter combined ignition-burning period
than that of petrol. This results in the revised ignition
timings being required to realise full power from an engine
converted from petrol to propane operation.
The LPG gives higher rates of cylinder pressure rise during
combustion indicating a faster combustion rate than that of petrol.
The LPG generally gives higher peak pressures than those of
petrol. However, because of the steeper rate of pressure
increase with LPG and the increase in the negative work
(compression before TDC) , the overall cycle work under the
pressure-time curve is slightly less than that of petrol.
This is reflected in the slightly lower torque values obtained
with LPG,
For variations in operating parameters (engine speed, ignition
timing and air fuel ratio) the two fuels are found to react
similarly. As engine speed increases, the peak pressures
decrease in magnitude and the peak occurs at a later crank
angle position.
22 -
As the point of ignition is advanced from top dead centre,
the peak pressures increase in magnitude for both fuels; the
UPGr developing slightly higher peak pressures as a result of
the shorter combined ignition-burning period.
Results show that, for ignition timings further advanced from
the optimum, the engine output is more sensitive with LPG,
whereas, for ignition timings retarded from the optimum, the
engine output is more sensitive with petrol. This is explained
by the differences between the combined ignition-burning
periods of the two fuels and the effect on the areas under the
pressure-time curves.
Variation of the air fuel ratios shows that the maximum rate
of pressure rise and peak pressure occurs,for both fuels,
in the region of leanest mixture for best torque. A leaner or
richer mixture results in a decrease in the rate of pressure
rise and a decrease in the peak pressures which occur at a
later angle.
The particular techniques used for these experiments, ie,
photographic recording of single traces of the pressure-time
traces from an oscilliscope screen, are limited in the sense
that proper statistical evaluation of many samples of the cyclic
variations that occur in cylinder pressure development could
not be made.
- 23 -
A better understanding of events taking place during combustion
in the spark ignition engine is now being obtained with modern
techniques such as the use of ionisation gap sensors from which
combustion rate data is monitored and statistically a n a ly s e d by10computer. Investigations by such people as Curry and Starkman
11et al are helping to obtain such an understanding.
- 24 -
BIBLIOGRAPHY
1. BUREAU OF TRANSPORT ECONOMICS. Liquified Petroleum Gas as a
Motor Vehicle Fuel. Australian Government Publishing Service. Canberra, 1974.
2. BYRNE, M.G. A Comparison of the Performance and Exhaust
Emissions of a Multi-Cylinder Engine Using Petrol and
Propane as Fuels. B.E. Thesis submitted to Wollongong University College, 1972.
3« BONAMY, S.E. and BYRNE, M.G. A Comparison of the Performance
and Exhaust Emissons of a Multi-Cylinder Engine Using Petrol
and Propane as Fuels. Thermofluids Conference 1974, Melbourne. The Institution of Engineers, Australia.
4. OBERTjE. Internal Combustion Engines. International Textbook Co., Penn. 19 6 8 .
5. LICHTY, L. Internal Combustion Engines. 6th edition,
McGraw-Hill, N.Y., 1951.
6. PATTERSON, D.J. Cylinder Pressure Variations, a Fundamental
Combustion Problem. SAE Transactions, Vol. 75> 1967» Paper
No. 660129.
7. RASSWEILER, G.M.,WITHROW, L. and CORNELIUS, W. Engine
Combustion and Pressure Development. SAE Transactions 46,
Jan. 1940.
- 25 -
8. MARVIN, C.F. and BEST, R.D. Flame Movement and Pressure
Development in an Engine Cylinder. NACA Report No.399, 1931.
9. MARVIN, C.F.,WHARTON,A. and ROEDER, C.H. Further Studies
-Of Flame Movement and Pressure Development in an Engine Cylinder NACA Report No.556,1936.
10. CURRY, S. A Three-Dimensional Study of Flame Propagation in a
_Spark Ignition Engine. SAE Transactions, Vol.71,1963. p638-650.
11. STARKMAN,E .S., STRANGE,F.M. and DAHM,T.J. Flame Speeds and
Pressure Rise Rates in Snark Ignition Engines. SAE Paper 83V, Sept.1959.
12. ADAMS,W. and B0LDT,K. What Engines Say About Propane Fuel Mixtures. SAE Transactions, Vol.73, 1965.
13• TAYLOR,C.*F• The Internal Combustion Engine in Theory and Practice. Vol.l, 1968. Wiley,N.Y.
14, BRITISH STANDARDS INSTITUTION. Methods for the Measurement
of Fluid Flow in Pipes. Part 1,B.S. 1042, 1964.
15* ALLSUP, J.R. and FLEMING, R.D. Emission Characteristics of
Propane as Automotive Fuel. Report of Investigations 7672,
Bureau of Mines, U.S. Dept, of the Interior.
16. BROWN, W.L. Methods for Evaluating Requirements and Error
in Cylinder Pressure Measurement. SAE Transactions, Vol.76,
Paper N0.67OOO8.
- 26 -
17* McCULLOUGH, J.D. Engine Cylinder Pressure Measurements,
SAE Transactions, Vol.6l, 1953.
18. GALSTERjG.M., GARNER,D.A. and BUCKLEY,E.D. What Propane
Engines Say About Spark Plugs. SAE Transactions, Vol. 74,1966.
19* BLALOCK,W.S. and LITTLE,R.N. Carburetion and Other Factors
which Affect Propane Fueled Engines. SAE Publication 285,
1966. Paper No.670058.
20. STANSFIELD, R. and WITHERS,J.G. Pressure Indicating for
Internal Combustion Engines. The Engineer, Aug. 1948.
21. WITHERS, J.G. Effects of Indicator Passages on the Accuracy
of Indicator Diagrams. The Engineer,Dec. 1955«
SECTION V
APPENDICES
- 27 -
APPENDIX A
EQUATIONS FOR ENGINE PERFORMANCE FACTORS. ’
POWER« For the particular dynamometer used the power is given
by the equation stated by the manufacturers,
EBP = W x N (l)400
where: BHP = brake horse power.
W = dynamometer balance reading.
N = engine speed, rpm.
TORQUE« The relationship between power and torque is-
BHP = 27TNT (2)33,000
where T = torque, ft lbf.
Equating (l) and (2) gives an expression for torque in terms of W-
T = 13.13 W (3)FUFT. CONSUMPTION. For petrol,the fuel consumption in lbm/hr.
is calculated from the time (t) required to use 100 ccs of
petrol of specific gravity 0.78.
lbm/hr = 100 x 35.515 x 10 ^x 62.4 x 0.78 x 3600t
= 618.79 (4)t
For the propane, the fuel consumption was measured in grams over
a period of 60 seconds. Thus the consumption in lbm/hr is given
by- lbm/hr = m x 60 = 0.1322 m (5)433.6
^ere: m = grams/60 secs.
- 28 -
AIR FUEL RATIO.
A/F ratio = mass flow rate of air (6)mass flow rate of fuel
The mass flow rate of air is obtained from the equation for the
orifice plate given at the end of Appendix B, The mass flow rate
of fuel is obtained from equation (4) or (5).
VOLUMETRIC EFFICIENCY, The volumetric efficiency ) is given
for a particular engine speed as
/! - act x 100Mth
where: ^act= ac^ua ̂mass flow rate of air.= theoretical mass flow rate to fill the piston
displacement volume at atmospheric conditions.
For the engine used, having a swept volume of 2268 cc and
suming the density of the air as 0.0747 lb/ft^ the equationas
becomes-
where
^ V =Mact x 100
N0.1797 N
engine speed, rpm.
(7)
- 29 -
APPENDIX B
ORIFICE PLATE CALCULATIONS USING
B,S, 1042 Part I.14
Assuming 80$ volumetric efficiency of the engine it was found
that the mass flow rate of air required at the maximum speed was 575 lbm/hr.
It is desired to have a pressure drop of approximately 2 inches
(50 mm) H^O across the orifice at the maximum flow rate.
An orifice plate with corner tappings (Section 7 of the Code)
is chosen as it is suited for measuring flow from a large space
into a pipe, ,
The downstream pipe has an average internal diameter of 6,46 inches.
According to Clause 49a of the Code, the orifice diameter can not
exceed half the downstream pipe diameter.
The atmospheric conditions for the initial design are taken as-
Temperature, T= 20°C (68°F)
Pressure, P=14.696 lhf/in
Relative humidity, $= 70$
2
- 30 -
To find the approximate orifice diameter required, the equations 10 and 12a of Clause 14a are used.
and
where:
N= W359.2 D2 h2j>2
CmE = NZ£
V = mass flow rate. (lhm/hr.)
D = downstream pipe dia. (ins)
h = differential across orifice, (ins H^O)
f = density of air. (lhm/ft^)
C = basic coefficient..,2
m = area ratio =D‘
(10)
(12a)
d = orifice dia. (ins)
E = velocity of approach factor.
Z = correction factor for pipe diameter and Reynolds No. £ = expansibility factor.
To calculate the density at the assumed conditions, Clause 25c
gives- p = 2.700 S (P - p ) + 0.62KT V T
where: <£ = specific gravity = 1.000 for air.
K = gas law deviation coeff. = 0.999 (Fig.6)
T = 528°R
P = 14.696 lhf/in2
p - partial pressure of water vapour = 0 p^ ̂
ft = 0.7
Pvs= 0.339 (Table 5)Substitution of these values gives the density as 0.07473 Tbm/ft .
- 31 -
With substitution equation 10 becomes-
N = 0.0993
To find an approximate value of mE it is assumed that Z and & in equation 12a are unity and thus-
CmE = N = 0.0993
and for this application of flow from a large area into a pipe,
Clause 56a states that the basic coefficient C = O.596. Thus-mE = N = 0.166
CAppendix J of the Code gives for mE = 0.166 the ratio
Therefore-d_D
=0.404
d = (0.404)(6.46) = 2.61 ins.
For convenience, an orifice diameter of 2.500 inches will be considered.
The more exact calculations for the mass flow rate equation will
now be carried out for a 2.500 inch diameter orifice. Equation 7
of Clause 13 of the Code gives the mass flow rate as-0 A iW = 359.2 CZ£E d h2j>2 (7)
and Rd = W15.8/cd (7)
where: W = mass flow rate of air (lbm/hr)
d = orifice diameter (ins)
C = basic coeff.
Z = Z Z r dZ = Reynolds number correction factor.
Z^= pipe diameter correction factor.
- 32 -
£ = expansibility factor.
E = velocity of approach factor,
h = pressure differential across orifice (ins H20)
p - air density at working conditions,
ylt = absolute viscosity of air (poise)Rd = Reynolds number.
Now have d = 2.500ins, D = 6.46 ins, C = 0.596(Clause 56e) and
for the assumed atmospheric conditions find /a= 1.81 x 10_4poise from Fig. 18. Therefore-
RR = ________ 575_______ = 80,425 (Eqn.7)1.81 x 10“4x 2.500
Zr = 1.000 for Rd = 80,000 (Fig.35b)
= 1.000 (Clause 56g)
Taking the specific heat ratio as X = 1.4 (Fig.21)
and for a maximum pressure drop of h = 2 ins H^O at a pressure
of P = 14.696 lbf/in^ giving —~ = O.I36, Figure 36 gives the
expansibility factor as £ = 0.998.
For —— = 0.387» Appendix J gives the velocity of approach factor
as E = 1.0115.j
The density was found to be jo = 0.07473 lbni/fx at the assumed
conditions.
When these values are substituted into Equation 7, the mass flow
rate equation becomes-1
W = 369.412 h2 where h = inches H^O1
W = 73.299 h *2 where h*= mm H20.or
- 33 -
Checking the differential pressure for the estimated maximum
flow rate gives h* = 6l,4mm H^O. This is an acceptable differential
and hence the 2.500 inch diameter is chosen.
The orifice plate and pressure tappings were made in accordance
with the specifications of Clauses 54 and 55 of the Code. The
minimum allowable length of downstream pipeline (Clause 48) for
the area ratio of m = 0.150 is five times the inside diameter of
the pipe, ie, 32.3 ins. The actual length used was 36-̂ ins. thus satisfying the condition.
In actual operation, the volumetric efficiency of the engine at
the maximum speed was less than the estimated 80$ and the
maximum pressure drops were in the order of 50 mm H^O , the
original desired maximum.
Thus, for the calculation of mass flow rates of air to the engine
the equation becomes- ^
W = 268.007 2.700 (P - Py)0.999 T
0.62 p+ _____fVT
1h »2
j
where: ¥ = mass flow rate (lbm/hr)
P = atmospheric pressure, (lbf/in )
T = atmospheric temperature. (°R)
p = partial pressure of water vapour in the air.
h®= differential pressure across orifice, (mm H 0)
- 34 -
APPENDIX C
TABULATION OF TEST DATA
TableCombined Ignition-Burning Periods and Maximum
Rates of Pressure Rise ±
Torque versus Ignition Advance - 2000 rpm 2
Torque versus Ignition Advance - 2500 rpm 3
Torque versus Ignition Advance - 3000 rpm 4
Torque versus Air Fuel Ratio - 2000 rpm 5
Torque versus Air Fuel Ratio - 2500 rpm 6
-35 -
TABLE 1
COMBINED IGNITION-BURNING PERIODS AND
MAXIMUM RATES OE PRESSURE INCREASE
SPEED FUEL TIMING TORQUE IGNITION- MAX RATE OF
BURNING PERIOD PRESSURE
INCREASErpm °BTDC ft.Ibf. msec (deg) psi/msec
(psi/deg)
2000 propane 28 95 4.25 (51) 108 (9)petrol 32 100 5.25 (63) 72 (6)
2500 propane 29 94 4.26 (64) 135 (9)petrol 32 99 4.80 (72) 90 (6)
3000 propane 29 88 3.56 (64) 144 (8)petrol 4.0 (72) 144 (8)32 95
PROPANE
PETROL
TABLE 2
TORQUE VERSUS IGNITION ADVANCE
2000 RPM FULL THROTTLE
- 3 6 -
Air fuel ratio 15 :1Ambient temperature 23°CAtmospheric pressure 759.4 mm Hg
IGNITION ADVANCE TORQUE°BTDC ft.lbf2 73.5
10 88.020 93.328 94.940 91.850 83.1
Air fuel ratio 12 : 1Ambient temperature 21°CAtmospheric pressure 762.4 mm Hg
IGNITION ADVANCE TORQUE°BTDC ft.lbf3 74.6
10 85.520 95.632 98.940 97.650 94.3
- 37 -
PROPANE
PETROL
TABLE 5
TORQUE VERSUS IGNITION ADVANCE
2500 RPM FULL TBROTTLE
Air fuel ratio 14.8 î 1Ambient temperature 23°CAtmospheric pressure 759*4 mm Hg
IGNITION ADVANCE TORQUE°BTDC ft.lbf2 71.9
10 84.720 92.329 93.640 91.450 85.1
Air fuel ratio 12 s 1Ambient temperature 23°CAtmospheric pressure 762.0 mm Hg
IGNITION ADVANCE TORQUE°BTDC ft.lbf2 69.310 86.420 96.332 100.240 98.950 94.6
- 38 -
PROPANE
PETROL
TABLE 4
TORQUE VERSUS IGNITION ADVANCE 3000 RPM FULL THROTTLE
Air fuel ratio 14.4 î 1Ambient temperature 23°CAtmospheric pressure 759.2 mm Hg
IGNITION ADVANCE TORQUEBTDC ft.lbf2 63.010 76.820 85.629 87.340 84.450 79.4
Air fuel ratio 12.5 s 1Ambient temperature 23°CAtmospheric pressure 762.0 mm Hg
IGNITION ADVANCE TORQUE°BTDC ft.lbf
3 65.710 79.420 90.232 92.741 90.3
- 39 -
PROPANE
PETROL
TABLE 5
TORQUE VERSUS AIR FUEL RATIO
2000 RPM FULL THROTTLE
Ignition advance 28°Ambient temperature 24°C Atmospheric pressure 759«0 mm Hg
AIR FUEL EQUIVALENCE TORQUERATIO RATIO ft.lbf13.9 0.886 92.214.1 0.896 92.414.4 0.92 92.415.3 0.972 92.517.5 1.111 86.418.1 1.15 76.823.4 1.49 32.2
Ignition advance 32°Ambient temperature 22°CAtmospheric pressure 760.0 mm Hg
AIR FUEL EQUIVALENCE TORQUERATIO RATIO ft.lbf9.6 0.645 97.49.9 0.662 98.510.3 0.69 99.311.0 0.735 99.811.8 0.79 99.312.6 0.847 97.614.9 1.00 88.017.2 1.158 67.018.0 1.21 55.2
- 40 - TABLE 6
PROPANE
PETROL
TORQUE VERSUS AIR FUEL RATIO
2500 RPM FULL TBROTTLE
Ignition advance 28°Ambient temperature 24°C Atmospheric pressure 759*0 mm Hg
AIR FUEL RATIO 12.2
12.8
13.0
13*9 17*2 17.5 18.1
Ignition advance Ambient temperature
EQUIVALENCE TORQUE RATIO ft.lbf0.775 91*10.813 91*30.83 91*50.885 91*61.094 85.41.12 65.71.15 34.2
32°22°C
Atmospheric pressure 760.0 mm Hg
AIR FUEL EQUIVALENCE TORQUERATIO RATIO ft.lbf9*9 0.666 97*210.3 0.69 98.711.1 0.746 98.911.9 0.795 98.612.6 0.845 97*714.8 0.995 89.316 .2 1.09 73.517.3 1 .16 57.7
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